High-flow-capacity centrifugal hydrogen gas compression systems, methods and components therefor

ABSTRACT

Hydrogen gas compression systems that each include a multistage centrifugal compressor in which each stage has an inlet-to-outlet pressure rise ratio of about 1.20 or greater. In one embodiment, the multistage compressor includes six high-speed centrifugal compressors driven at a speed of about 60,000 rpm. The compressor has an output of more than 200,000 kg/day at a pressure of more than 1,000 psig. The compressors for the compression stages are distributed on both sides of a common gearbox, which has gearing that allows axial thrusts from the compressors to be handled effectively. Each stage&#39;s compressor has a unique impeller, which is secured to a support shaft using a tension-rod-based attachment system. In another embodiment, the multistage compressor is driven by a combustion turbine and one or more intercoolers are provided between compression stages. Each intercooler is cooled by coolant from an absorption chiller utilizing exhaust gas from the combustion turbine.

RELATED APPLICATION DATA

This application is a divisional of U.S. Nonprovisional patentapplication Ser. No. 13/259,934, filed on Sep. 23, 2011, and titled“High-Flow-Capacity Centrifugal Hydrogen Gas Compression Systems,Methods and Components Therefor”, which is a U.S. National Phase ofPCT/US2010/028449, filed Mar. 24, 2010, that claims the benefit ofpriority of U.S. Provisional Patent Application Ser. No. 61/162,753,filed Mar. 24, 2009, and titled “Centrifugal Hydrogen Gas CompressionSystem And Method.” Each of these applications is incorporated byreference herein in its entirety.

FIELD OF THE INVENTION

The present invention generally relates to the field of gas compression.In particular, the present invention is directed to high-flow-capacitycentrifugal hydrogen gas compression systems, methods and componentstherefor.

BACKGROUND

The growth and establishment of a hydrogen economy requires an efficientand cost effective hydrogen compressor for high volumetric transport ofhydrogen through pipelines. The U.S. Department of Energy's NationalHydrogen Energy Roadmap reported, “the primary challenge to using morehydrogen in our energy systems is the cost of producing, storing, andtransporting hydrogen.” A practical hydrogen economy requires a hugeinfrastructure buildup, including production, storage and delivery, toreplace the currently existing gasoline and natural gas infrastructure.

Significant research is ongoing in the area of very high-pressurehydrogen compression for refueling station applications in whichpressures up to 12,000 psi are desired. These systems includemulti-stage reciprocating piston and diaphragm compressors,liquid-piston compressors, electro-chemical compression (high-pressureelectrolyzers), and metal and chemical hydride gas compression. However,for hydrogen pipeline applications, the flow rates and powerrequirements are many orders of magnitude greater, and as such,reciprocating gas compressor technology is currently the dominantapproach. Current hydrogen pipeline compressors capable of high volumesare capital intensive with poor reliability and high maintenance cost.To meet the future needs of the hydrogen infrastructure, advanced highefficiency compressors that overcome these issues are required.

SUMMARY OF THE DISCLOSURE

In one implementation, the present disclosure is directed to a system.The system includes at least one centrifugal compressor including: animpeller having a rotational axis and including an impeller body havinga blind hole extending along the rotational axis only partway into theimpeller body; an impeller support shaft configured to support theimpeller, the support shaft having a first end, a second end spaced fromthe first end, and a central bore extending from the first end to thesecond end; and a tension rod having a first end disposed in the blindhole and non-rotatably coupled to the impeller, the tension rod beingdisposed in the central bore and configured to secure the impeller tothe impeller support shaft.

In another implementation, the present disclosure is directed to ahydrogen gas compression system designed for remote operation. Thesystem includes a hydrogen gas compressor including a plurality ofcentrifugal compressor stages, each of the plurality of centrifugalcompressor stages having a working fluid inlet and working fluid outlet;at least one intercooler for cooling a hydrogen gas working fluid priorto the hydrogen gas working fluid entering at least one of thecentrifugal compressor working fluid inlets; a turbine operativelycoupled to the hydrogen gas compressor for driving the plurality ofcentrifugal compressor stages, wherein the turbine utilizes the hydrogengas compressed by the hydrogen gas compressor as fuel for combustion andproduces an exhaust during operation; and an absorption chilleroperatively connected to the at least one intercooler for providing acoolant thereto, wherein the absorption chiller utilizes the turbineexhaust to cool the coolant.

In yet another implementation, the present disclosure is directed to amethod of providing high pressure hydrogen for transport. The methodincludes providing a hydrogen gas compressor including a plurality ofcentrifugal compressor stages each having an aluminum-alloy impellerwith forward-swept trailing edges; compressing hydrogen with thehydrogen gas compressor; and providing hydrogen compressed during thecompression step for transport.

BRIEF DESCRIPTION OF THE DRAWINGS

For the purpose of illustrating the invention, the drawings show aspectsof one or more embodiments of the invention. However, it should beunderstood that the present invention is not limited to the precisearrangements and instrumentalities shown in the drawings, wherein:

FIG. 1 is a high-level schematic diagram of a high-flow-capacitycentrifugal hydrogen gas compression system made in accordance with thepresent disclosure;

FIG. 2 is an isometric view of a six-stage multistage compressor thatcan be used with the system of FIG. 1;

FIG. 3A is a partial cross-sectional/partial side view of animpeller/shaft assembly that can be used in each of the six stages ofthe multistage compressor of FIG. 2;

FIG. 3B is a cross-sectional view of an alternative impeller/shaftassembly that can be used in each of the six stages of the multistagecompressor of FIG. 2;

FIG. 4 is a front view of each of the impellers of the correspondingrespective impeller shaft/assemblies of FIGS. 3A-B;

FIG. 5 is a cross-sectional view of the gearbox of FIG. 2 as taken alongline 5-5 of that figure;

FIG. 6 is a cross-sectional view of the gearbox of FIG. 2 as taken alongline 6-6 of that figure;

FIG. 7 is a cross-sectional view of the gearbox of FIG. 2 as taken alongline 7-7 of that figure;

FIG. 8 is a high-level schematic diagram of an alternative centrifugalhydrogen gas compressor system that utilizes a combustion turbine as theprime mover for the multi-stage compressor;

FIG. 9 is an exemplary performance map for a particular example of themultistage compressor shown in FIG. 2;

FIG. 10 is a schematic diagram illustrating a model used to formulate athermo-fluid analytic model for use in selecting a proper surge valvethat inhibits surge during an emergency shutdown of a centrifugalcompressor;

FIG. 11 is screenshot of the user interface of software used forselecting a proper surge valve that inhibits surge during an emergencyshutdown (ESD) of a centrifugal compressor;

FIG. 12A is an exemplary performance map for a particular example of themultistage compressor shown in FIG. 2, showing the ESD shutdown path ofthe compressor when the surge-control valve flow coefficient Cv value is50 ft³/sec/(psig)^(0.5);

FIG. 12B is an exemplary performance map for the multistage compressorcorresponding to FIG. 12A, showing the ESD shutdown path of thecompressor when the surge-control valve flow coefficient Cv value is 31ft³/sec/(psig)^(0.5);

FIG. 12C is an exemplary performance map for the multistage compressorcorresponding to FIGS. 12A-B, showing the ESD shutdown path of thecompressor when the surge-control valve flow coefficient Cv value is 30ft³/sec/(psig)^(0.5);

FIG. 12D is an exemplary performance map for the multistage compressorcorresponding to FIGS. 12A-C, showing the ESD shutdown path of thecompressor when the surge-control valve flow coefficient Cv value is 38ft³/sec/(psig)^(0.5);

FIG. 13 is a flow diagram illustrating a method of selecting asurge-control valve for a gas compression system; and

FIG. 14 is a high-level schematic diagram illustrating a computingdevice representative of computing devices that can be used inimplementing the method of FIG. 13.

DETAILED DESCRIPTION

Referring now to the drawings, FIG. 1 illustrates a high-flow-capacitycentrifugal hydrogen gas compression system 100 that has a number ofunique features that allow it to compress hydrogen gas to a dischargepressure greater than 1,000 psig at a flow rate of greater than 200,000kg/day with 98% efficiency, while maintaining 99.99% hydrogen purity andall in a relatively small package size. As those skilled in the art willreadily appreciate, system 100 is highly suitable for application in ahydrogen distribution system of a hydrogen economy and is an economical,efficient replacement for reciprocating compressors in such adistribution system. The unique features of system 100 include, but arenot limited to: 1) a special impeller design that enables very highspeeds; 2) an special impeller/shaft assembly design that readilyenables the use of differing materials; 3) a unique gearbox design thatbalances thrust loads and 4) a unique control system design that allowssystem 100 to operate relatively near its surge instability region so asto maximize performance. Each of these and other unique features ofsystem 100 is described below in detail.

In this example, system 100 is designed to provide a hydrogen dischargepressure of about 1,285 psig at a flow rate of about 240,000 kg/day andwith a hydrogen efficiency of 98% (btu/btu) and a hydrogen purity of99.99%. System 100 achieves this performance using six sequentiallyarranged compression stages 104A-F with five intercooling heatexchangers 108A-E fluidly coupled between adjacent ones of the stages.In this example, each compression stage 104A-F comprises a correspondingcentrifugal compressor 112A-F, which is driven by a prime mover 116 viaa common gearbox 120. The combination of compressors 112A-F and gearbox120 is referred to hereinafter as a “multistage compressor” based on thefunctionality of the combination. A particular example of a multistagecompressor 200 suitable for use as the multistage compressor of FIG. 1is described below in connection with FIGS. 2-7. Prime mover 116 can beone or more fixed-speed electric motors, variable-speed electric motors,steam turbines, combustion turbines, reciprocating internal combustionengines, or any combination thereof, among others. In one of theexamples below, prime mover 116 is a four-pole fixed-speed electricmotor having a rotor speed of 1,800 rpm. In another of the examplesbelow, prime mover 116 is a combustion turbine.

System 100 includes an inlet valve 124 and an outlet check valve 128. Inthis example, inlet valve 124 is fluidly coupled to a hydrogen source(not shown) that provides hydrogen at a mass flow rate of about 22,000lb/hr at a pressure of about 350 psi and a temperature of about 100° F.Outlet check valve 128 is fluidly coupled to, for example, a pipelinesystem (not shown) for carrying the hydrogen to one or more destinationlocations. When system 100 is operating normally, outlet check valve 128outputs compressed hydrogen at a mass flow rate of about 21,900 lb/hr ata pressure of about 1,295 psi and a temperature of about 145° F. Abypass/surge-control system 132, which includes valve 132A andassociated piping 132B, is provided for bypassing compression stages104A-F and for controlling the operation of system 100 during startup,during transient conditions, such as emergency shutdown, and so that thesystem operates at peak performance during normal operation.Bypass/surge control system 132 may also include a surge control heatexchanger 132C to reduce the suction temperature of recirculatinghydrogen in the event that continuous recirculation is needed to preventsurge. In this example, heat exchanger 132C is plumbed in parallel withheat exchangers 108A-E to a supply 136 and return 140 of a suitablecoolant, such as water.

In one example wherein the pressures and mass flow rates are asmentioned above, each of compressors 112A-F is driven to a rotationalspeed of about 60,000 rpm. When prime mover 116 is an electric motorhaving a fixed rotor speed of 1,800 rpm, gearbox 120 must, therefore,provide a final drive ratio of about 33:1. If prime mover 116 is adevice operating at a higher speed, such as a steam turbine, combustionturbine or a variable speed electric motor driven at a higher frequency,the final drive ratio that gearbox 120 must provide can be much lower.

As mentioned above, FIG. 2 illustrates a particular six-stage multistagecompressor 200 that can be used as the multistage compressor incentrifugal hydrogen gas compression system 100 of FIG. 1. As seen inFIG. 2, multistage compressor 200 includes a gearbox 204 and sixcompressors 208A-F (only compressors 208A-E are seen in FIG. 2) thatprovide six sequential stages of compression, such as compression stages104A-F of FIG. 1. It is noted that the piping necessary for sequentiallycoupling compressors 208A-F to one another and/or to intercoolers is notshown for the sake of clarity. Each compressor 208A-F includes acorresponding compressor housing 212A-F (only housings 212A-D are seenin FIG. 2) that houses a corresponding impeller (not show, but may beany suitable impeller, such as either one of impellers 304, 348 of FIGS.3A-B, respectively). As will be described below in greater detail, eachof impellers 304, 348 is 8 inches (20.32 cm) in diameter and driven at apeak speed of 60,000 rpm during normal operation when used in thesix-stage example of FIG. 1 with the mass-flow and pressure conditionsnoted above.

Gearbox 204 includes a housing 216 having a front face 216A and a rearface 216B spaced from one another to accommodate gearing and otherinternal components (see FIGS. 5-7). In this context, the terms “front”and “rear” are used simply to designate the locations of the facesrelative to the prime mover (not shown) that provides power to gearbox204 via an input shaft 220. In this example, three compressor housings212A-C are mounted to gearbox 204 on front face 216A and threecompressor housings are mounted to the gearbox on rear face 216B. Thisdesign, in combination with gearing (FIGS. 5-7) inside gearbox 204,provides a number of advantages, including compactness of multistagecompressor 200 and handling of thrust loads on the gearing fromcompressors 208A-F.

FIG. 3A illustrates an impeller/shaft assembly 300 that can be used ineach of compressors 208A-F of FIG. 2. Impeller/shaft assembly 300includes an impeller 304, a shaft 308 and an impeller attachment system312 that secures the impeller to the shaft. Impeller attachment system312 provides an important benefit of readily permitting the use ofdiffering materials for impeller 304 and shaft 308 so as to optimizethose materials without compromising on compressor performance. In oneexample, impeller 304 is made of a high-strength aluminum alloy, such asa 7075-T6 or 2618-T7 alloy, or other metal, such as a high-strengthtitanium alloy coated with a coating that protects the titanium alloyfrom the hydrogen being compressed. While impeller 304 can be made ofany material suitable for the conditions of compressing hydrogen gas,shaft 308 can be made of a durable metal typically used forturbomachinery shafts, such as steel. Impeller attachment system 312also provides an important benefit of being able to make the rotor body304A without a full-length central bore, thereby avoiding stressconcentrators in high-stress regions of the rotor body. This allowsimpeller 304 to be operated at the high speeds necessary to achieve thehigh performance of a multistage hydrogen gas compressor, such ascompressor 200 of FIG. 2.

In this example, shaft 308 has a stepped design, which includes a sealregion 308A for receiving a suitable seal 316 (which would form a sealwith gearbox housing 216 of FIG. 2), two radial/axial bearing regions308B-C for receiving, respectively, suitable radial/axial bearings320A-B, and a pinion region 308D containing teeth 324 that engagegearing within gearbox 204 (FIG. 2). In one example, teeth 324 have ahelical configuration, as discussed below in more detail. Shaft 308 alsoincludes a central bore 328 extending the full length of the shaft andcontaining a tension rod 332 of impeller attachment system 312 thatsecures impeller 304 to the shaft.

In addition to tension rod 332, impeller attachment system 312 includesa threaded insert 336 engaged with impeller 304 in a central bore 340 inthe impeller. Insert 336 can be secured to impeller 304 in any suitablemanner, such as by threaded engagement or using a heat-shrink fit.Insert 336 threadedly receives a corresponding threaded end 332A oftension rod 332. In one example, threaded insert 336 is made of steel,though another suitable material can be used. In this embodiment, insert336 and the impeller-engaging end of shaft 308 include inter-engagingface splining 344 to inhibit rotation between the two parts and betweenimpeller 304 and the shaft. It is noted that in other embodiments, suchas the embodiment shown in FIG. 3B, an insert such as insert 336 of FIG.3A need not be used. For example, in FIG. 3B, tension rod 348 directlythreadedly engages a threaded central bore 352 in impeller 356.

Returning to FIG. 3A, impeller attachment system 312 of this examplealso includes nut 360 that threadedly engages tension rod 336. Aretainer 364 and a set 368 of spring washers is provided to assist inensuring that nut 360 is not over-tightened. In other embodiments, suchas the embodiment shown in FIG. 3B, such a spring-washer/retainerarrangement is not used, and other means are implemented for ensuringthe correct amount of tension is induced into tension rod 348. As thoseskilled in the art will readily appreciate, when tension rod 332 of FIG.3A and tension rod 348 of FIG. 3B are properly tensioned to secure therespective impeller 304, 356 to the respective shaft 308, 352, theshafts are placed into compression to counteract the tension in thetension rods. The use of impeller attachment system 312 provides anumber of advantages. For example, face splining 344 can be used insteadof a simply spin-on type arrangement between impeller 304 and shaft 308,thereby providing a much more positive engagement of the impeller withthe shaft. In addition, with impeller attachment system 312 impeller 304can be readily removed, for example during maintenance outages, withoutremoving shaft 308.

FIG. 4 illustrates a blade configuration 400 that can be used on each ofimpellers 304, 356 of FIGS. 3A-B, respectively. In this example, therotation of each impeller 304, 356 during use is counterclockwise in thedirection of arrow 404. In this example, blade configuration 400utilizes a dual split-impeller design having eight long blades 408A-H,eight intermediate-length blades 412A-H and sixteen short blades 416A-Parranged so that each intermediate-length blade is located midwaybetween a corresponding pair of long blades and each short blade islocated midway between a corresponding long blade and an adjacentintermediate-length blade. Each of blades 408A-H, 412A-H, 416A-P isforward-swept at each of the inlet and exit regions 400A-B of bladeconfiguration 400. In other words, both the leading and trailing ends ofblades 408A-H, 412A-H, 416A-P, relative to the flow of hydrogen duringoperation, is curved in the direction of rotation direction 404. In theexample shown, the trailing ends of blades 408A-H, 412A-H, 416A-P have aforward sweep of about 30° in the direction of rotation. This is theangle of the blades at the trailing edge as measured from a radial lineextending from the axis of rotation.

The incorporation of forward curved blades 408A-H, 412A-H, 416A-P atexit region 400B provides an increased pressure ratio, and the dualsplit impeller design provides higher efficiency and pressure ratio. Theuse of these and other improvements to conventional process compressorimpeller designs has resulted in improvements in impeller efficiencywhile also maintaining a high impeller stage pressure ratio thatadvances the art for high capacity hydrogen compression. The forwardsweep of blades at exit region 400B increases the exit tangentialvelocity for a higher work coefficient. The reduced blade count at inletregion 400A resulting from the split impeller design reduces inletblockage for higher pressure ratio and efficiency at a given flow rate,and the second set of splitter blades 416A-P towards exit region 400Breduces the blade loadings, which increase as the inlet blade countdecreases. These factors contribute to the high efficiency of bladeconfiguration 400 as well as an increase in the work input andachievable pressure ratio.

In this example, the outside diameter, Do, of blade configuration 400,and hence each impeller 304, 356, is 8 inches. When each impeller 304,356 is driven at a speed of 60,000 rpm, the stresses within thatimpeller are safely handled by the material used, for example, any ofthe alloys mentioned above and the corresponding hydrogen gascompression system, such as system 100 of FIG. 1, can provide very goodperformance. As an example, the following Table provides performancedetails for a case in which the inlet pressure and temperature of thehydrogen provided to system 100 (FIG. 1) at inlet valve 124 are 350 psiand 100° C., respectively.

TABLE Example Performance Details Stage 1 Stage 2 Stage 3 Stage 4 Stage5 Stage 6 Overall Speed (rpm)  60000      60000     60000      60000    60000      60000    Power (HP)  1388     1388    1449     1469   1427    1445   8566 Eff TT Var P    80.6%    80.8%    82.3%    81.6%   81.2%     80.4% Pressure Ratio TT    1.25    1.25    1.27    1.27   1.26     1.26 Specific Speed    0.64    0.58    0.49    0.44    0.40    0.36 Specific Speed (US)  82.5  74.4  63.7  56.8  52.1    46.5 U2(ft/s) 2094    2094    2094    2094    2094    2094   Impeller Diam (in) 8  8  8  8 8   8 Backsweep (B2B) (deg)  0  0  0  0 −20   −20  T02M (F)146.2 146.0 147.9 148.3 146.7  146.9 Pexit (psig) 438.1 535.8 667.9834.2 1035.7  1287.4 Flow Coeffs Phi      0.178      0.146      0.119     0.095     0.077       0.062 Misc Work Coefficient      0.916     0.917      0.957      0.970     0.942       0.954

Referring now to FIGS. 5-7, and also to FIGS. 2 and 3A, gearbox 204 isdesigned for a final drive ratio of 33.3:1 so that when input shaft 220is driven at a speed of 1,800 rpm, each of compressors 208A-F is drivento 60,000 rpm by the gearbox. Gearbox 204 achieves this final driveratio using a large bull gear 500 that is directly driven by input shaft220. Large bull gear 500 meshes with three smaller pinion gears 504A-Cthat each drive a pair of small bull gears 508A-F, one on each side ofthe large bull gear, at a speed ratio of 10. Each small bull gear 508A-Fdrives a corresponding one of compressors 208A-F (FIG. 2) at a speedratio of 3.33, for a final drive ratio of 33.3:1. As will be appreciatedby those skilled in the art, when shaft 308 of FIG. 3A is used for eachof compressors 208A-F, each small bull gear 508A-F meshes with suchshaft at pinion region 308D. An important benefit to using thedual-bull-gear arrangement of gearbox 204 in which pairs of compressors208A-F on opposite faces 216A-B are driven by a correspondingdual-bull-gear assembly 512A-C is that thrust loads from the compressorsare balanced. As mentioned above, pinion region 308D on shaft 308 ofimpeller/shaft assembly 300 has helical teeth 324. Each small bull gear508A-F has matching helically-configured teeth 514. Because of thehelical configuration of meshing teeth 324, 514, thrust loads fromcompressors 208A-F on opposing faces of gearbox 204 driven by aparticular common dual-bull-gear assembly 512A-C are transmitted to thatassembly. Because such opposing compressors 208A-F are driven in thesame direction on opposing sides of gearbox 204, during normal operationthe axial thrusts therefrom are in opposing directions so that theygenerally cancel one another within that dual-bull-gear assembly 512A-C.

FIG. 8 illustrates a hydrogen gas compression system 800 that can besimilar to or the same as system 100 of FIG. 1, except that the primemover for powering the multistage compressor 804 is a combustion turbine808. Since hydrogen gas is being compressed by system 800, combustionturbine 808 can use some of the process gas as the fuel for combustion.This makes system 800 particularly suitable for remote locations wherepower and other fuels are not readily accessible. An important featureof system 800 is the inclusion of an absorption chiller system 812 thatuses heat from the exhaust 816 of combustion turbine 808 to chill acoolant 820, such as water, for use in one or more heat exchangers 822(only one shown for convenience) for cooling the hydrogen process gas atany one or more stages of the compression process, such as prior tocompression, between sequential stages, after compression and inrecirculation, and any combination thereof. Absorption chiller system812 includes one or more absorption chillers 824 (only one shown forconvenience) that may be any suitable type of chiller, including onethat uses ammonia as the refrigerant and water as the absorbent and onethat uses ammonia, hydrogen gas and water. Various types of absorptionchillers suitable for use in absorption chiller system 812 are wellknown in the art such that it is not necessary to describe them in anydetail herein for those skilled in the art to understand how to make anduse hydrogen gas compression system 800.

In this example, multistage compressor 804 is a four-stage compressorhaving four centrifugal compressors 828A-D driven by a common gearbox832, which in turn is driven by combustion turbine 808. Each compressor828A-D may be the same as or similar to each compressor 112A-F of FIG.1, and consequently can include either of impeller/shaft assemblies 300,372 of FIGS. 3A-B. In other embodiments, compressors 828A-D can haveconfigurations different from the configurations described aboverelative to FIGS. 1-4 and/or can be provided in a different number thanshown in FIG. 8. As mentioned above, multistage compressor 804 can bereplaced with multistage compressor 200 of FIG. 2, if desired. In thatcase, however, gearbox 204 will have to be modified to change the finaldrive ratio if the output speed of combustion turbine 808 is not 1,800rpm and it is desired to drive compressors 828A-D to the same 60,000 rpmspeed discussed above.

Combustion turbine 808 can be any combustion turbine having thenecessary output-shaft power for driving at least multistage compressor804. Preferably, however, combustion turbine 808 is configured for usinghydrogen from the process gas as the fuel. Optionally, one or moreauxiliary electrical power generators 836 also driven by combustionturbine 808 can be provided for generating electrical power. Theelectrical power generated by generator(s) 836 can be used, for example,to power the electronics that form part of hydrogen gas compressionsystem 800, such as the control electronics. Of course, the generatedelectrical power can be used for other purposes as well.

Combustion turbine 808 includes a compressor 840, for compressing air,and one or more combustors 844, where the fuel is added to thecompressed air in order to heat the air. The heated exhaust fromcombustor(s) 844 is then passed through an expansion turbine 848, wherethe exhaust from the combustor(s) is expanded to provide the powernecessary to drive combustion-turbine compressor 840, multistagehydrogen gas compressor 804, and any auxiliary generator(s) 836 or othermachinery (not shown). The exhaust from expansion turbine 848 isprovided to absorption chiller system 812 to provide the heat needed topower the absorption cycle for chilling coolant 820. In this embodiment,combustion turbine 808 also includes a recouperator 852 that uses heatin the exhaust from expansion turbine to pre-heat the air provided tocombustor(s) 844 from combustion-turbine compressor 840.

To achieve the high performance of a hydrogen gas compression system ofthe present disclosure, the multistage compressor is operated relativelyvery close to instability. If the system is not designed correctly, anexcursion, such as an emergency shutdown (ESD), can cause the compressorto surge. As those skilled in the art know, during surge a compressorloses its ability to maintain peak head and the entire system becomesunstable and experiences flow reversal. Surge causes high fluctuatingloads on the compressor's thrust bearings and can result in catastrophicdamage to the compressor if left unabated. FIG. 9 illustrates aperformance map 900 for the six-stage compressor system 100 describedabove. In this example, performance map 900 approximates the surge line904 as being linear and shows three performance lines, line 908 showingthe performance at 60,000 rpm, line 912 showing the performance at about50,000 rpm and line 916 showing the performance at about 40,000 rpm. Innormal operation, multistage compressor 200 (FIG. 2) operates to theright of surge line 904 with a margin of 5% to 10% and surge-to-riseratios of 6% minimum considered acceptable. In one example, theoperating point of multistage compressor 204 is 6.1 lbm/s. As seen fromline 908 on performance map 900, at this mass flow rate the overallcompression ratio is about 3.7.

Referring to FIG. 1, during normal operation, bypass/surge-controlsystem 132 is used to control adequate flow rate through compressors112A-F by recirculating compressed gas from last stage 104F back to theinlet of first stage 104A. An intercooler, such as heat exchanger 132C,is often needed to reduce the suction temperature if continuousrecirculation is required to prevent surge. Operating system 100 veryclose to its optimum head vs. flow rate operating point requires veryprecise and rapidly responsive instrumentations and valves.

ESD conditions also require very precise and rapidly responsiveinstrumentations and valves. During an ESD, the rotational mass inertiaof prime mover 116, for example, an electric motor, will continue todrive compressors 112A-F. During the ESD sequence, the flow rate throughcompressors 112A-F may be reduced to less than the surge limit if therecirculated flow rate cannot relieve the pressure between the outlet oflast compressor stage 104F and outlet check valve 128 fast enough. Inorder to ensure that system 100 is provided with a suitably sized andsuitably fast-acting surge valve, the present inventors developed asurge valve design method that can be used to select a surge valve basedon ESD conditions. As described below in detail, this method utilizes athermo-fluid analytical model based on the model 1000 shown in FIG. 10.

Referring to FIG. 10, model 1000 includes a compressor 1004 having aninput 1004A and an output 1004B, discharge piping 1008, a surge-controlvalve 1012, an outlet check valve 1016 and the volume 1020 of gastrapped in the discharge piping between the compressor and each of thesurge-control and outlet check valves. The stored energy from the massmoment of inertia of the compressor (e.g., multistage compressor 200 ofFIG. 2) and prime mover 116 (FIG. 1) will continue to compress hydrogenuntil the stored energy of the compressor/prime mover system isdissipated with time. The rate of change of energy (compression power)for the system is given by the following Equation 1a as a function ofthe instantaneous speed, N, and the rate of change of speed,

$\left( \frac{\partial N}{\partial t} \right).$

$\begin{matrix}{{{Compressor}\mspace{14mu}{Power}} = {\left( \frac{I_{{motor} + {compressor}}}{g_{c}} \right) \times N \times \frac{\partial N}{\partial t}}} & \left\{ {{{Eq}.\mspace{14mu} 1}a} \right\}\end{matrix}$Wherein:

I_(motor+compressor) is the mass moment of inertia of the motor (primemover) and compressor; and

g_(c) is the gravitational constant=32.2 Lbm-ft/Lbf/s².

The complete equation with the expression for the compression power as afunction of volume flow rate and pressure ratio is given by Equation 1b,below.

$\begin{matrix}{{\overset{.}{Q} \times \rho \times {Cp} \times {Tr} \times \left\lfloor {\left( \frac{Po}{Pi} \right)^{\frac{k - 1}{k}} - 1} \right\rfloor} = {I \times \frac{4\pi^{2}}{g_{c}} \times N \times \frac{\partial N}{\partial t}}} & \left\{ {{{Eq}.\mspace{14mu} 1}b} \right\}\end{matrix}$Wherein:

{dot over (Q)} is the volume flow rate of the hydrogen gas;

ρ is the density of the hydrogen gas;

Cp is the specific heat of the gas in BTU/Lbm/F;

Tr is the compressor inlet's absolute temperature (Rankine);

Po is the outlet pressure of the compressor;

Pi is the inlet pressure of the compressor;

k is the specific heat ratio (Cp/Cv);

I is the mass moment of inertia of the motor (prime mover) andcompressor; and

g_(c) is the gravitational constant=32.2 Lbm-ft/Lbf/s².

As the stored rotational energy of the compressor/primer-mover systemcontinues to compress the hydrogen gas, the trapped volume 1020 of gaswill be relieved if a surge valve having a sufficiently large flowcoefficient Cv is used, and compressor surge can be prevented if the gasis recirculated to the suction of the compressor. The flow coefficientCv can be determined from the following Equation 2.{dot over (Q)}=C _(v)×(P _(n) −P _(suction))^(1/2)  {Eq. 2}Wherein:

{dot over (Q)} is the volume flow rate of the hydrogen gas;

P_(n) is the nth pressure increment out of the compressor; and

P_(suction) is the suction pressure of the compressor.

The equation that determines the amount of hydrogen gas that can passthrough the surge valve (e.g., valve 132A in FIG. 1) is provided inEquation 3, below.

$\begin{matrix}{{\left( {\overset{.}{M_{in}} - \overset{.}{M_{out}}} \right) \times R_{u} \times \frac{T}{V \times {MoleWt}}} = \frac{\partial P}{\partial t}} & \left\{ {{Eq}.\mspace{14mu} 3} \right\}\end{matrix}$Wherein:

{dot over (M)}_(in) is the mass flow rate of the hydrogen gas into thevalve;

{dot over (M)}_(out) is the mass flow rate of the hydrogen gas out ofthe valve;

R_(u) is the universal gas constant=1,545 ft-Lbf/Lbmole/R;

T is the absolute temperature at the discharge of the compressor(Rankine);

V is the volume capacity of the discharge piping (ft³);

MoleWt is mole weight of the hydrogen gas; and

∂P/∂T is the rate of change of pressure with respect to time.

The rate of change of the pressure for the hydrogen gas trapped involume 1020 is based on the formulation shown in Equation 3 and is afunction of the amount of gas that is being delivered to the trappedvolume by compressor 1004 and the amount of gas that is being relievedvia surge-control valve 1012.

FIG. 11 illustrates a screenshot 1100 of a graphical user interface(GUI) 1104 of software (not shown) that mathematically models model 1000of FIG. 10 (graphically represented for the benefit of the user in thescreenshot at 1108). GUI 1104 allows a user to readily input into thesoftware values for all of the fixed parameters that will be used foranalyzing the impact of differing valve coefficients on the operation ofthe subject compressor, for example, multistage compressor 200 of FIG.2, during an ESD event. As mentioned above, a goal of this analysis isto inform a designer in selecting a surge-control valve, such as valve132A in FIG. 1, that is large enough, but also fast enough, to preventthe compressor from surging during the ESD event. It is noted that thesurge-control valve cannot be excessively large because this usuallyimplies a very large and costly valve. As seen in FIG. 11, in thisexample, input parameters include parameters 1112 for calculating themass moment of inertia of the gearbox and compressors of multistagecompressor 1004, parameters 1116 for modeling the multistage compressor,and parameters 1120 for modeling piping 1008 and its volume 1020 betweenthe compressor and outlet check valve 1016, among other parametersneeded for the analysis.

The calculations using the thermo-fluid analytical model of the softwarebegins with the design point operation of the pressure ratio and flowrate at time t=0 seconds and proceeds with determining the nextcompressor flow rate based on the incremental change in the compressorspeed using Equation 1b, above. The change in the pressure of thehydrogen gas trapped in volume 1020 determines the discharge pressure ofthe compressor during the next time step, here Δt=0.271 seconds, usingEquation 3, above. This new pressure is used in conjunction with theperformance curves (not shown) for compressor 1004 to determine the newflow rate that the compressor will deliver as a function of the newdischarge pressure at this next time step as determined by Equation 3.This new flow rate and discharge pressure then begins anothercalculation at a new time step, t_(n±1). The two initial calculationsare reflected in lines 1124, 1128, respectively, of GUI 1104.

As can be readily understood by those skilled in the art, once thephysical sizes of piping 1008 and trapped-gas volume 1020 and theparticular data for compressor 1004 and the prime mover are input intoGUI 1104, the only parameter that is changed in the analysis is themagnitude of the surge-control valve flow coefficient Cv 1132. Bychanging the value of flow coefficient Cv, the thermo-fluids modelcalculates the path 1200 of the pressure ratio versus flow rate forcompressor 1004 and can be plotted on a performance map 1204 asillustrated in FIG. 12A. A properly selected surge-control valve flowcoefficient (and, correspondingly, surge-control valve itself), such asthe one that yields path 1200 in FIG. 12A, will fall to the right ofsurge line 1208. As can be readily seen, path 1200 stays completely tothe right of surge line 1208 and, therefore, compressor 1004 does notbecome unstable during an ESD event. In this particular example, thesurge-control valve flow coefficient Cv has a value of 50ft³/sec/(psig)^(0.5).

A very large surge-control valve may require a measurable amount of timebefore it can respond to the ESD event that has been caused to thesystem by a number of different reasons. This time delay causes thepressure in trapped volume 1020 (FIG. 10) of the gas to increase with asubsequent decrease in the flow rate that can only be overcome with aneven larger surge valve. The thermo-fluids analytic method of thepresent disclosure determines the speed of sound in the trapped gas(e.g., hydrogen gas) and hence the time delay in the control's responseto a system ESD. However, the response time of the pressure andtemperature instruments and the processing time for the control systemthat is used with the surge control system must be known and used as aninput to the ESD surge model.

A small change in the surge-control valve coefficient Cv can have alarge effect on the safe performance of a compressor system. Forexample, FIGS. 12B-C, reveal performance paths 1224, 1228, respectively,when compressor 1004 that is in danger of imminent surge failure andultimate surge failure. In the case of FIG. 12B, the value ofsurge-control valve flow coefficient Cv used in the model is 31ft³/sec/(psig)^(0.5), and, in FIG. 12C, the value of surge-control valveflow coefficient Cv used in the model is 30 ft³/sec/(psig)^(0.5). Theseare to be compared with ESD path 1232 shown in FIG. 12D where thesurge-control valve flow coefficient Cv is 38 ft³/sec/(psig)^(0.5).

With the foregoing in mind, FIG. 13 illustrates an exemplary method 1300of selecting a properly sized surge-control valve for an ESD event in acentrifugal gas compression system in which residual rotational energyfrom the mass inertia of rotating parts within the compressor and primemover driving the compressor keeps driving the compressor after power orfuel to the prime mover is shutdown. At step 1305, a computer model ofthe gas compression system, such as model 1000 of FIG. 10, is created.This step involves inputting into the model, for example, using a GUIthe same as or similar to GUI 1104 of FIG. 11, values for variousparameters needed to run the model, such as parameters 1108, 1112, 1116,1120 of FIG. 11. At step 1310, an initial value of the flow coefficientCv for the surge-control valve is input into the model, and, at step1315, the computer model is run for various increments of time followingthe time when the providing of power/fuel for the prime mover is cut andthe moving parts of the prime mover and compressor begin to wind down.At this step, the computer model performs the iterative calculationsdescribed above.

At step 1320, an ESD path is established, typically, by the computerused to run the computer model. Examples of the ESD path generated atstep 1320 include paths 1200, 1224, 1228, 1232 graphically shown,respectively, in FIGS. 12A-D. The ESD path established at step 1320 canbe in any suitable form for further processing, such as a collection ofoverall-compression-pressure-ratio-versus-mass-flow-rate data points. Inother embodiments, the data points calculated at step 1315 can besubjected to a curve-fitting algorithm and/or plotted on a performancemap for graphical analysis.

At step 1325, the ESD path is compared to a surge line for thecompressor at issue to determine its relation to the surge line. Thiscan be accomplished using automated algorithms for determining where thecalculated data points of the ESD path or any point on a curve fittedthereto are located relative to the surge line. At step 1325, thelocation of the ESD path is assessed to determine whether or not itsatisfies one or more criteria established for a properly sizedsurge-control valve. For example, the criteria may be that the ESD pathbe to the right of the surge line and that the closest point on the ESDpath be spaced at least a certain minimum distance (in properperformance map unit terms) from the surge line. This minimum distancemay include a safety margin for modeling errors and other factors ofsafety. On the other hand, it might be determined that the ESD path istoo far to the right of the surge line and, therefore, that the flowcoefficient Cv is too large such that the resulting physicalsurge-control valve would be too large and/or too slow-acting. Step 1325can be performed automatically by the computer or may be performed by adesigner, for example, viewing a plot of the ESD path on a performancemap.

At step 1330 it is determined whether or not the criterion(ia) at step1325 is/are satisfied. If the criterion(ia) is/are determined to besatisfied at step 1330, method 1300 proceeds to step 1335 wherein thecurrent value of the flow coefficient Cv is used to select asurge-control valve. If, however, at step 1330 it is determined that thecriterion(ia) at step 1325 is/are not met, method 1300 proceeds to step1340 wherein a new value for the flow coefficient Cv is selected.Depending on how the ESD path failed the criterion(ia), the new valueselected can be either larger or smaller than the value just used. Forexample, if the determination at step 1330 was that the ESD path was tooclose to the surge line, then at step 1340 a larger flow-coefficientvalue is selected. On the other hand, if the determination at step 1330was that the ESD path was too far to the right of the surge line, thenat step 1340 a smaller flow-coefficient value is selected. Once a newvalue for flow coefficient Cv has been selected at step 1340, method1300 proceeds to step 1315 at which point the model is re-run with thenew value. Steps 1315 through 1340 are repeated until the ESD pathsatisfies the desired criterion(ia), allowing a proper surge-controlvalve to be selected at step 1335.

FIG. 14 shows a diagrammatic representation of one embodiment of acomputing device in the exemplary form of a computer system 1400 withinwhich a set of instructions for causing the device to perform any one ormore of the aspects and/or methodologies of the present disclosure maybe executed. Computer system 1400 includes a processor 1404 (e.g., amicroprocessor) (more than one may be provided) and a memory 1408 thatcommunicate with each other, and with other components, via a bus 1412.Bus 1412 may include any of several types of bus structures including,but not limited to, a memory bus, a memory controller, a peripheral bus,a local bus, and any combination thereof, using any of a variety of busarchitectures well known in the art.

Memory 1408 may include various components including, but not limitedto, a random access read/write memory component (e.g, a static RAM(SRAM), a dynamic RAM (DRAM), etc.), a read-only component, and anycombination thereof. In one example, a basic input/output system 1416(BIOS), including basic routines that help to transfer informationbetween elements within computer system 1400, such as during start-up,may be stored in memory 1408.

Memory 1408 may also include (e.g., stored on one or moremachine-readable media) instructions (e.g., software) 1420 embodying anyone or more of the aspects and/or methodologies of the presentdisclosure. In another example, memory 1408 may further include anynumber of instruction sets including, but not limited to, an operatingsystem, one or more application programs, other program modules, programdata, and any combination thereof.

Computer system 1400 may also include one or more storage devices 1424.Examples of storage devices suitable for use as any one of the storagedevices 1424 include, but are not limited to, a hard disk drive devicethat reads from and/or writes to a hard disk, a magnetic disk drivedevice that reads from and/or writes to a removable magnetic disk, anoptical disk drive device that reads from and/or writes to an opticalmedium (e.g., a CD, a DVD, etc.), a solid-state memory device, and anycombination thereof. Each storage device 1424 may be connected to bus1412 by an appropriate interface (not shown). Example interfacesinclude, but are not limited to, Small Computer Systems Interface(SCSI), advanced technology attachment (ATA), serial ATA, universalserial bus (USB), IEEE 13144 (FIREWIRE), and any combination thereof. Inone example, storage device 1424 may be removably interfaced withcomputer system 1400 (e.g., via an external port connector (not shown)).Particularly, storage device 1424 and an associated machine-readablestorage medium 1428 may provide nonvolatile and/or volatile storage ofmachine-readable instructions, data structures, program modules, and/orother data and/or data storage for computer system 1400. In one example,instructions 1420 may reside, completely or partially, withinmachine-readable storage medium 1428. In another example, instructions1420 may reside, completely or partially, within processor 1404.

In some embodiments, such as a general purpose computer, computer system1400 may also include one or more input devices 1432. In one example, auser of computer system 1400 may enter commands and/or other informationinto the computer system via one or more of the input devices 1432.Examples of input devices that can be used as any one of input devices1432 include, but are not limited to, an alpha-numeric input device(e.g., a keyboard), a pointing device, a joystick, an audio input device(e.g., a microphone, a voice response system, etc.), a cursor controldevice (e.g., a mouse), a touchpad, an optical scanner, a video capturedevice (e.g., a still camera, a video camera), touchscreen, a digitizerpad, and any combination thereof. Each input device 1432 may beinterfaced to bus 1412 via any of a variety of interfaces (not shown)including, but not limited to, a serial interface, a parallel interface,a game port, a Universal Serial Bus (USB) interface, a FIREWIREinterface, a direct interface to the bus, a wireless interface (e.g., aBluetooth® connection) and any combination thereof.

Commands and/or other information may be input to computer system 1400via storage device 1424 (e.g., a removable disk drive, a flash drive,etc.) and/or one or more network interface devices 1436. A networkinterface device, such as network interface device 1436, may be utilizedfor connecting computer system 1400 to one or more of a variety ofnetworks, such as network 1440, and one or more remote devices 1444connected thereto. Examples of a network interface device include, butare not limited to, a network interface card, a modem, a wirelesstransceiver (e.g., a Bluetooth® transceiver) and any combinationthereof. Examples of a network include, but are not limited to, a widearea network (e.g., the Internet, an enterprise network), a local areanetwork (e.g., a network associated with an office, a building, acampus, a group of wireless sensors or other group of data streamingdevices, or other relatively small geographic space), a telephonenetwork, a direct connection between two computing devices, and anycombination thereof. A network, such as network 1440, may employ a wiredand/or a wireless mode of communication. In general, any networktopology may be used. Information (e.g., data, instructions 1420, etc.)may be communicated to and/or from computer system 1400 via the one ormore network interface devices 1436.

In some embodiments, such as a general purpose computer, computer system1400 may further include a video display adapter 1448 for communicatinga displayable image to a display device, such as display device 1452.Examples of a display device include, but are not limited to, a liquidcrystal display (LCD), a cathode ray tube (CRT), a plasma display, andany combination thereof. In addition to a display device, a computersystem 1400 may include one or more other peripheral output devicesincluding, but not limited to, an audio speaker, a printer, and anycombination thereof. Such peripheral output devices may be connected tobus 1412 via a peripheral interface 1456. Examples of a peripheralinterface include, but are not limited to, a serial port, a USBconnection, a FIREWIRE connection, a parallel connection, and anycombination thereof.

A digitizer (not shown) and an accompanying pen/stylus, if needed, maybe included in order to digitally capture freehand input. A pendigitizer may be separately configured or coextensive with a displayarea of display device 1452. Accordingly, a digitizer may be integratedwith display device 1452, or may exist as a separate device overlayingor otherwise appended to the display device.

Exemplary embodiments have been disclosed above and illustrated in theaccompanying drawings. It will be understood by those skilled in the artthat various changes, omissions and additions may be made to that whichis specifically disclosed herein without departing from the spirit andscope of the present invention.

What is claimed is:
 1. A system, comprising: a hydrogen gas compressor including a plurality of centrifugal compressor stages, each of said plurality of centrifugal compressor stages having a working fluid inlet and working fluid outlet; at least one intercooler for cooling a hydrogen gas working fluid prior to the hydrogen gas working fluid entering at least one of the centrifugal compressor stages working fluid inlets; a turbine operatively coupled to said hydrogen gas compressor for driving said plurality of centrifugal compressor stages, wherein said turbine produces an exhaust during operation; and an absorption chiller operatively connected to said at least one intercooler for providing a coolant thereto, wherein said absorption chiller utilizes the turbine exhaust to cool said coolant wherein said turbine includes a combustor that utilizes hydrogen gas compressed by said hydrogen gas compressor for combustion, said system being configured for operation in remote locations without access to energy sources other than the hydrogen gas working fluid.
 2. A system according to claim 1, wherein at least one of said centrifugal compressor stages includes: an impeller having a rotational axis and including an impeller body having a blind hole extending along said rotational axis only partway into said impeller body; an impeller support shaft configured to support said impeller, said support shaft having a first end, a second end spaced from said first end, and a central bore extending from said first end to said second end; and a tension rod having a first end disposed in said blind hole and non-rotatably coupled to said impeller, said tension rod being disposed in said central bore and configured to secure said impeller to said impeller support shaft.
 3. A system according to claim 2, wherein said first end of said impeller support shaft and said impeller each include complementary features that allow relative movement between said impeller and said impeller support shaft along said rotational axis and prevent relative rotational movement, said tension rod configured to maintain said complementary features in engagement.
 4. A system according to claim 3, wherein said at least one centrifugal compressor further includes an insert disposed in said blind hole, said tension rod coupled to said insert, further wherein said insert includes a face that confronts said face of said first end of said impeller support shaft, said face of said insert including splining designed and configured to engage said splining on said face of said impeller support shaft.
 5. A system according to claim 2, wherein said tension rod has a second end, said at least one centrifugal compressor further including an adjustment means for coupling said second end of said tension rod to said second end of said impeller support shaft and for adjusting a tension of said tension rod.
 6. A system according to claim 2, further comprising: a hydrogen gas compressor including: a plurality of said at least one centrifugal compressor configured as a plurality of centrifugal compressor stages; a gearbox for transmitting a rotational force to said plurality of centrifugal compressor stages, said gearbox including a first gear and a plurality of dual-gear assemblies engaging said first gear so as to be driven thereby, wherein each of said plurality of dual-gear assemblies includes second and third gears for driving corresponding respective ones of said plurality of centrifugal compressor stages.
 7. A system according to claim 6, wherein said impeller support shaft of each of said plurality of centrifugal compressor stages includes a pinion region containing teeth that engage a corresponding respective one of said second or third gears of one of said plurality of dual-gear assemblies.
 8. A system according to claim 7, wherein said gearbox includes a housing having first and second opposed sides, said plurality of centrifugal compressor stages including a first set of said centrifugal compressor stages disposed on said first side and a second set of said centrifugal compressor stages disposed on said second side so that, during use, an axial thrust generated by each of said first set of said plurality of centrifugal compressor stages is substantially cancelled by an axial thrust generated by a corresponding respective one of said second set of said plurality of centrifugal compressor stages.
 9. A system according to claim 6, wherein at least one of said impellers of said plurality of centrifugal compressor stages has an outlet region and a plurality of blades, wherein at least one of said plurality of blades are forward-swept at said outlet region.
 10. A system according to claim 9, wherein each of said impellers of said plurality of centrifugal compressor stages is formed from an aluminum alloy.
 11. A system according to claim 1, wherein: the plurality of centrifugal compressor stages include impellers, wherein at least one of said impellers of said plurality of centrifugal compressor stages has an outlet region and a plurality of blades, wherein at least one of said plurality of blades is forward-swept at said outlet region.
 12. A system according to claim 11, wherein said plurality of blades includes a plurality of splitter blades.
 13. A hydrogen gas compression system designed for remote operation, comprising: a hydrogen gas compressor including a plurality of centrifugal compressor stages, each of said plurality of centrifugal compressor stages having a working fluid inlet and working fluid outlet; at least one intercooler for cooling a hydrogen gas working fluid prior to the hydrogen gas working fluid entering at least one of the centrifugal compressor working fluid inlets; a turbine operatively coupled to said hydrogen gas compressor for driving said plurality of centrifugal compressor stages, wherein said turbine includes a combustor that utilizes the hydrogen gas compressed by the hydrogen gas compressor as fuel for combustion and produces an exhaust during operation; and an absorption chiller operatively connected to said at least one intercooler for providing a coolant thereto, wherein said absorption chiller utilizes the turbine exhaust to cool said coolant.
 14. A hydrogen gas compression system according to claim 13, wherein each of said plurality of centrifugal compressor stages includes an impeller, at least one of said impellers having an outlet region and a plurality of blades, wherein at least one of said plurality of blades is forward-swept at said outlet region.
 15. A hydrogen gas compression system according to claim 14, wherein each of said impellers is formed from an aluminum alloy.
 16. A hydrogen gas compression system according to claim 13, wherein each of said plurality of centrifugal compressor stages includes an impeller and an impeller support shaft supporting a corresponding said impeller, wherein a tension rod secures each said impeller to a corresponding said impeller support shaft.
 17. A method of providing high pressure hydrogen for transport, comprising: providing a hydrogen gas compressor including a plurality of centrifugal compressor stages each having an aluminum-alloy impeller with forward-swept trailing edges; compressing hydrogen with the hydrogen gas compressor; providing hydrogen compressed during the compression step for transport; and providing a portion of the hydrogen compressed during the compressing step as fuel for powering a turbine configured to power the hydrogen gas compressor.
 18. A method according to claim 17, wherein the plurality of aluminum-alloy impellers are integrally geared with a common gearbox to form a compact design.
 19. A method according to claim 17, further including: providing a portion of an exhaust from the turbine to an absorption chiller for cooling a coolant; providing the coolant to an intercooler for cooling hydrogen compressed by one of the centrifugal compressor stages before the hydrogen is further compressed by another one of the centrifugal compressor stages.
 20. A method according to claim 17, wherein said compressing step includes operating each impeller between approximately 50,000 rpm and approximately 60,000 rpm to provide approximately 200,000 kg/day to approximately 240,000 kg/day at a discharge pressure of approximately 1,000 psig to approximately 1,285 psig. 